Hydraulic valve actuation systems and methods

ABSTRACT

Hydraulic engine valve actuation systems and methods for internal combustion engines. The systems utilize a proportional valve to regulate the flow of a working fluid to and from a hydraulic actuator controlling the engine valve position. The position of the proportional valve is controlled by high speed valves to control various engine valve parameters, including engine valve takeoff and landing velocities. Returning all valves to a known starting position between engine valve events avoids accumulation of errors in proportional valve positioning. Embodiments using spool valves for the high speed valves and the proportional valve, and spring return and hydraulic return for the engine valve, are disclosed. A specially shaped spool in the proportional valve provides enhanced control over the engine valve operation. Various further alternate embodiments are disclosed.

CROSS-REFERENCE TO RELATED APPLICATION

[0001] This application is a continuation of application Ser. No.09/729,487, filed Dec. 4, 2000, entitled “Hydraulic Valve ActuationSystems and Methods.”

BACKGROUND OF THE INVENTION

[0002] 1. Field of the Invention

[0003] The present invention relates to the field of hydraulic valveactuation for internal combustion engines.

[0004] 2. Prior Art

[0005] At the present time, piston-type internal combustion engines ofinterest to the present invention are currently widely used inautomobiles, trucks, buses and various other mobile and stationary powersystems. Such engines include the common gasoline and diesel engines, aswell as similar engines operating from different fuels such as liquidpropane. These engines commonly utilize intake and exhaust valves thatare spring loaded to the closed position and which are directly orindirectly opened at appropriate times by a camshaft driven from theengine crankshaft. In a two-cycle engine such as a two-cycle dieselengine, the camshaft will rotate in synchronism with the enginecrankshaft, though in a four-cycle engine, the camshaft is driventhrough a two-to-one reduction drive system (gear or chain or belt,etc.) to rotate at one-half the engine crankshaft speed.

[0006] Camshaft actuation of engine valves historically has had a numberof advantages, resulting in its relatively universal use in such enginesfor many decades. These advantages include high reliability,particularly given the current level of development of such cam actuatedvalve systems. Cam actuation is also relatively cost effective, againgiven the state of development and quantities in which it is produced.Cam actuation also has the advantage of allowing shaping the cam toprovide a smooth curve defining valve position versus camshaft angle.This results in a rather low velocity takeoff and initial valve opening,as well as a rather low velocity valve final closing at low enginespeeds, resulting in minimum noise being generated. It also results infaster valve opening and valve closing at higher engine speeds asrequired to maintain the same valve timing throughout the engine speedoperating range.

[0007] Cam actuated valve systems also have certain limitations whichare becoming of increasing concern. In particular, optimal valve timingis not fixed throughout the engine operating range. For instance, valvetiming for maximum power at one engine speed will not be the same asvalve timing for maximum power at another engine speed. Accordingly, theclassic cam operated valve systems utilize a compromise valve timing,providing reasonable performance over a reasonable range of engineoperating conditions while being less than optimal for most, if not atall, these conditions. Further, valve timing for maximum power at anyengine speed may not be optimal from an engine emissions standpoint.Optimum valve timing at any given engine speed may need to be dependenton engine loading, and perhaps other parameters, such as airtemperature, air pressure, engine temperature, etc.

[0008] Recently, mechanisms have been introduced to attempt to make upfor some of the limitations in the fixed timing cam operated valvesystems. These mechanisms include mechanisms for varying valve timing(but not valve opening duration in terms of camshaft angle) with enginespeed, as well as mechanisms for also increasing the valve openduration. However, such mechanisms tend to be complicated, open thevalve a fixed distance under all engine operating speeds and are limitedin the number and range of variables for which valve operation may beginto be optimized.

[0009] Recently various hydraulic systems for valve actuation have beenproposed. These systems offer the potential of more flexible control ofvalve actuation parameters over the range of the various engineoperating parameters. The present invention is an improvement on thesesystems.

BRIEF SUMMARY OF THE INVENTION

[0010] Hydraulic engine valve actuation systems and methods for internalcombustion engines. The systems utilize a proportional valve to regulatethe flow of a working fluid to and from a hydraulic actuator controllingthe engine valve position. The position of the proportional valve iscontrolled by high speed valves to control various engine valveparameters, including engine valve takeoff and landing velocities.Returning all valves to a known starting position between engine valveevents avoids accumulation of errors in proportional valve positioning.Embodiments using spool valves for the high speed valves and theproportional valve, and spring return and hydraulic return for theengine valve, are disclosed.

[0011] To provide enhanced control over the engine valve operation, aspecially shaped spool in the proportional valve may be used to shapethe flow areas versus spool position. This allows more gradualrestricting of the flow areas versus spool movement over selectedportions of the possible spool positions, diminishing the effect ofsmall errors in spool position in such regions without inhibiting themaximum flow areas when the spool is at its maximum positions.

[0012] Various further alternate embodiments are disclosed.

BRIEF DESCRIPTION OF THE DRAWINGS

[0013]FIG. 1 is a block diagram of an exemplary configuration of asystem in accordance with the present invention.

[0014]FIG. 2 is a diagram illustrating the general structure andfunction of the three-way proportional spool valve 24 of FIG. 1.

[0015]FIG. 3 is a perspective view of the spool 38 of the proportionalvalve of FIG. 2.

[0016]FIG. 4 is an expanded view of an edge of the center land of thespool 38 of FIG. 3.

[0017]FIGS. 5 and 6 are graphical representations of the flow areaversus spool position provided by the proportional valve 24 between thehigh pressure rail and the chamber 26 of the valve actuator, and betweenthe chamber 26 and the vent 37, respectively.

[0018]FIG. 7 is a cross sectional view of an engine valve actuatorconsisting of two concentric pistons that may be used with the presentinvention.

[0019]FIG. 8 is a diagram of an embodiment of the present invention thatcontrols a hydraulically returned engine valve using a closed center3-way proportional valve.

[0020]FIG. 9 is a diagram of an embodiment of the present invention thatcontrols a hydraulically returned engine valve using a closed center4-way proportional valve.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

[0021] The present invention is a hydraulic valve operating system foroperating one or more intake valves or one or more exhaust valves in apiston-type internal combustion engine, which provides full flexibilityin valve timing, valve duration, extent of opening, and valve openingand closing velocity. Operation over the desired range of these andother parameters may be controlled, and more importantly optimized, forall engine operating conditions. Such optimization may also includeincrementally adjusting the valve operation based on the valve operationduring a previous valve operating cycle. This is achieved by controllingthe position of a proportional valve by the use of pilot valves tocontrol the operating parameters of an intake or exhaust valve. In thatregard, a reference herein and in the claims to an “intake valve” or an“exhaust valve,” unless otherwise made clear by the context in which thephrase is used, shall mean one or more intake valves for a cylinder ofan internal combustion engine, or one or more exhaust valves of acylinder of an internal combustion engine. Exemplary embodiments of thissystem, sometimes referred to herein as a “two-stage” system, arehereafter described in detail.

[0022] First referring to FIG. 1, a block diagram of an exemplaryconfiguration of a system in accordance with the present invention maybe seen. The system illustrated in FIG. 1 may be used to actuate anintake or an exhaust valve. This 2-stage system consists of 2 miniature2-way digital latching spool valves 20 and 22 coupled to control theposition of a 3-way proportional spool valve 24. The proportional spoolvalve, in turn, controls the flow area into, and out of, a controlvolume 26. This control volume acts on an actuator 28 to regulate theposition of the engine valve 30. In this embodiment, a spring return 32is utilized for valve closing, though embodiments with hydraulic valveclosing may also be used, as shall be subsequently described.

[0023] The 2 miniature 2-way digital latching spool valves 20 and 22(referred to herein as pilot valves) may preferably be identical valves,preferably in accordance with the 2 way valves disclosed in U.S. Pat.No. 5,640,987 entitled Digital Two, Three, and Four Way Solenoid ControlValves, issued Jun. 24, 1997, the disclosure of which is incorporatedherein by reference. Such valves are double solenoid, high speed,magnetically latching spool valves, that as used in the presentinvention, are operable between two positions. The first positioncouples a first port to a second port for fluid communication betweenthe two ports, and the second position blocks fluid communicationbetween the first and second ports. While other types of valves could beused, such as poppet valves, valves generally of the type disclosed inthe above referenced patent are preferred because of their very highspeed for good control, and low energy consumption because of suchcapabilities as their magnetic latching, and the ability to sensecompletion of actuation, if used, to minimize heating above the alreadyrelatively warm environment in which they operate. (See U.S. Pat. Nos.5,720,261 and 5,954,030.)

[0024] In the embodiment of the present invention of FIG. 1, valve 20allows fluid flow from fluid line 34 to a drain line or reservoir 37 (ata relatively low pressure, such as atmospheric pressure) when in itsfirst position, and blocks fluid flow from fluid line 34 to the drain 37when in its second position. Valve 22 allows fluid flow from a lowpressure rail 36 to the fluid line 34 when in its first position, andblocks fluid flow from the low pressure rail 36 to the fluid line 34when in its second position. Check valve 23 is optional, and is normallyclosed, as the differential pressure on the check valve normally willnot be in a direction to open the valve. Its presence however, will helpdamp transient pressure fluctuations and recover energy in the pressurefluctuations.

[0025] Now referring to FIG. 2, a diagram illustrating the generalstructure and function of the three-way proportional spool valve 24 ofFIG. 1 may be seen. The proportional spool valve includes a spool 38within an internal housing 40 which fits within an external housingassembly (not shown) with O-rings in O-ring grooves 42 to separate theregions of ports 1, 2 and 3 from each other and from the ends of theinternal housing 40. In that regard, the outer housing assembly, inaddition to having the associated fluid connections, also includesinternal annual grooves adjacent each of the regions identified as ports1, 2 and 3 in FIG. 2, each to act as a manifold region for the holesthrough the internal housing 40 for fluid communication with arespective one of the inner regions 44, 46 and 48 in the internalhousing 40, respectively. Fluid communication from each of the ports tothe associated inner region 44, 46 or 48 is provided in the exemplaryembodiment not only by through holes 50, but also by cooperativelydisposed orthogonal through holes 52 associated with each of the ports.

[0026] As schematically illustrated in FIG. 2, the spool 38 ispositioned within the internal housing 40 by fluid pressures acting on apiston at the left end of the spool having an effective area Al and apiston at the right side of the spool having an effective area of A₂. Asspecifically illustrated in FIG. 2, the spool 38 is shown in its extremeright position, referred to herein as its first position, as defined bystops on the travel of either the pistons actuating the spool or stopsacting on the spool itself. In this position, the spool 38 is blockingfluid communication between ports 3 and 2 and is allowing fluidcommunication between ports 2 and 1. Obviously, when the spool is at itsleft-most position, referred to herein as its second position, fluidcommunication between ports 1 and 2 is blocked and fluid communicationbetween ports 2 and 3 is enabled.

[0027] Normally in a spool valve, by way of example in the twominiature, two-way digital latching spool valves 20 and 22 of FIG. 1,fluid communication between two adjacent ports will be blocked when thespool is in one position and during the initial motion of the spooltoward the other position. However, once the relief on the spoolassociated with the land in the housing separating the regions coupledto the two adjacent ports starts to bridge the land, a flow area betweenthe regions coupled to the two ports is established, that flow areaincreasing linearly with further motion of the spool. Because that flowarea is a peripheral flow area of the full diameter of the spool, onceopening starts, a relatively large flow area between the two ports willbe opened with only a relatively small further motion of the spool.

[0028] However, in the three-way proportional spool valve 24 (FIG. 1),some of the details of which are illustrated in FIG. 2, this change inflow area versus spool position is purposely modified to reshape theflow area versus spool position. In the exemplary embodiment, this isaccomplished in the manner illustrated in FIGS. 3 and 4. In that regard,FIG. 3 is a perspective view of the spool 38 and FIG. 4 is an expandedview of an edge of the center land of the spool 38. As may be seen inFIG. 3, the center land on the spool has a plurality of kerfs 54 equallyspaced around each end of the center land, which kerfs begin to open acontrolled flow area with spool position prior to the edge of the landon the spool reaching the edge of the land on the internal housing, thenormal position for a spool valve flow area starting to be established.

[0029] In addition, as may be seen in FIG. 4, small steps are ground inthe center land of the spool of the three-way proportional spool valveof the exemplary embodiment. Thus, while the spool has an outer diameterDo having a close sliding fit within the inner diameter of the internalhousing, each end of the center land of the exemplary spool hasadditional diameters D₁, D₂ and D₃, where D₃ is less than D₂, D₂ is lessthan D₁ and D₁ is less than D₀. This provides a non-linear variation inflow area versus spool position during the opening and closing of thefluid communication between adjacent ports, as illustrated in FIGS. 5and 6. These figures illustrate the flow area between ports 1 and 2, andports 2 and 3, respectively, versus the position of the spool in thethree-way proportional spool valve. As may be seen in FIG. 5, when thespool is at the right-most position, the flow area between ports 1 and 2is a maximum, initially decreasing at a relatively high rate for theinitial motion of the spool to the right, then decreasing in rate foranother part of the motion, then decreasing at a further reduced rate toa substantially zero flow area for the rest of the spool motion,essentially blocking communication between ports 1 and 2 whenapproximately 40% of the spool motion has been achieved. In comparison,FIG. 6 shows the flow area between ports 2 and 3, which is a mirrorimage of FIG. 5.

[0030] It will be noted from FIGS. 5 and 6 that in the exemplaryembodiment of the present invention, the reduction in flow area on theinitial valve closing motion of the spool occurs at a high rate withrespect to spool position, decreasing in change in flow rate with anincreasing position of the spool until the flow area goes tosubstantially zero when less than half of the spool motion has beenachieved, thereby substantially altering the flow area versus spoolposition characteristic of a conventional spool valve. Also, because theflow area goes to substantially zero before one-half of the maximumspool travel has been achieved, fluid communication between both ports 1and 2, and ports 2 and 3, is disabled or blocked when the spool isapproximately centered within its travel range. For the specificexemplary embodiment illustrated, the substantial blockage between bothports 1 and 2, and ports 2 and 3, occurs whenever the spool's positionis anywhere between approximately 40% of its travel and 60% of itstravel. Obviously other shaping of the flow areas, or no shaping may beused if desired, though preferably some shaping will be used to diminishthe effect of small errors in spool position in the restricted regionswithout inhibiting the maximum flow areas when the spool is at itsmaximum positions.

[0031] Referring again to FIG. 1, it may be seen that fluid in the lowpressure rail 36, which may have a pressure, by way of example, of 20 to50 bar, is coupled to the right side of the three-way proportional spoolvalve 24 to act on the area A₂ (FIG. 2) of a piston encouraging thespool to its left-most position.

[0032] Assuming spool valve 22 is open and spool valve 20 is closed,pressure in the low pressure rail 36 is communicated to line 34, andthus acts on area A₁ of the piston actuating the spool of theproportional spool valve (FIGS. 1 and 2). Because the area A₁ is largerthan the area A₂, the spool of the proportional spool valve is forced toits right-most position, coupling port 1 and port 2 to couple chamber 26to vent, allowing the valve return spring 32 to force the valve 30 tothe closed position. Preferably area A₁ is approximately twice area A₂so that A₁−A₂ A₂.

[0033] If the two-way valve 22 is closed and the two-way valve 20 isopen, line 34 will be vented to the drain 37, so that the pressureacting on piston area A₁ (FIG. 2) of the three-way proportional spoolvalve will be substantially zero. The pressure acting on area A₂ of thespool valve, however, will be equal to the pressure of the low pressurerail 36, thereby creating an unbalanced force on the spool to force thespool to its left-most position. In this position, port 2 is in fluidcommunication with port 3, communicating the pressure in the highpressure rail 56 to control volume 26 to force the valve 30 open.

[0034] If, by way of example, valve 30 is half open and spool valves 20and 22 are both closed, then port 2 of the proportional spool valve willbe isolated from both ports 1 and 3, so that the fluid in the controlvolume 26 is trapped, maintaining the valve 30 at its present position.Finally, since the two-way spool valves 20 and 22 are very high speedvalves, they may be controlled in such as manner as to rapidlycontrollably place the spool of the proportional spool valve at anydesired location within the extremes of its travel, and thus variablycontrol the flow rate of fluid into or out of the control volume 26.This, in turn, allows full control of the operating parameters of thevalve 30, such as the extent of opening, the timing and duration ofopening, the velocity profile of the opening and closing of the valve(which profiles can be different from each other and/or vary with engineoperating conditions), and the final valve closing velocity with enginerpm. This allows a relatively low velocity valve closing at low enginerpm for low noise operation, while still allowing the closing velocityto be increased with engine rpm, as necessary for higher engineoperating speeds.

[0035] The fluid used in the exemplary embodiment in the low pressurerail, the high pressure rail and passed to drain is engine operatingoil, though other fluids may be used if desired. Since the flow rates inthe control system for valve 30 will vary with various parameters, suchas oil viscosity, and thus oil temperature, and the pressure of the lowpressure rail and the high pressure rail, operation of the valve controlsystem of FIG. 1 must reasonably compensate for such variations. As afirst order approximation, these variations may be reasonably modeled sothat the control system as shown in FIG. 1 can reasonably vary operatingdurations of valves 20 and 22 to at least approximate the desiredprofile of the proportional valve spool position with engine crankshaftangle, given the existing engine operating parameters (speed, engineload, fuel temperature, air temperature, engine oil temperature,atmospheric pressure, etc.).

[0036] In the exemplary embodiment, a small Hall effect sensor 58 ispositioned adjacent actuator 28 for the valve 30 so as to provide afeedback signal to the controller. Thus valve motion during a valveoperating cycle may be monitored and used to control the operation ofthe valves 20 and 22 for that valve operating cycle, and/or to makecorrections in the next valve operating cycle to more accurately achieveoptimum valve operation for that valve operating cycle. In that regard,more optimum operation may be determined in any of various ways,including better compliance to a predetermined valve position profileversus engine crank angle as predetermined for the then existing engineoperating conditions and ambient conditions, or as determined by theeffect of incremental changes on one or more engine performancecharacteristics for the change in valve operation just made, or acombination of both.

[0037] In the event two (or more) valves are being actuated in unison bya single proportional valve 24, a sensor such as a position sensor (Halleffect sensor or other position sensor) may be used on only one of thevalves, or on both valves, the sum of the signals providing a betteraverage indication of the position profile of the two valves and thedifference in the signals providing fault detection, such as a stickyvalve. While a position sensor(s) is preferred, other types of sensorscould be used, such as a velocity sensor, as the integration times toconvert to position are short. In that regard, at the end of each valveoperating cycle, the control valve 22 is actuated to couple line 34 tothe low pressure rail 36 and control valve 20 is actuated to decoupleline 34 from the drain 37 to bring the spool 38 to the stop at theposition shown schematically in FIG. 1. This provides predeterminedspool and pilot valve starting points for each valve operating cycle sothat errors in the spool valve position do not accumulate, valveoperating cycle to valve operating cycle. If desired, a sensor may alsobe used to sense the position of the proportional spool valve spool 38,though this is not preferred.

[0038] Thus the two miniature latching valves 20 and 22 (sometimesreferred to herein as pilot valves) control the position of theproportional valve 24. Specifically, the supply pilot valve 22 allowsfluid to flow between a low-pressure rail 32 (approximately 20-50 bar)and a first piston used to move the proportional 3-way valve. The ventpilot valve 20 will allow fluid to flow from the piston to a vent atatmospheric pressure. Using these pilot valves, the position of theproportional valve can be changed quickly and accurately. The positionof the proportional valve can be infinitely varied throughout 3 flowstates noted in FIGS. 5 and 6, namely:

[0039] State 1: The high pressure fluid from the high pressure rail 56(approximately 100-240 bar) is allowed to flow from the high pressurerail to a control volume 26 above the engine valve actuation piston.

[0040] State 2: The spool 38 of the proportional valve is centeredbetween its hard stops, trapping fluid in the control volume above theengine valve actuation piston and creating a hydraulic lock.

[0041] State 3: The fluid in the control volume 26 above the enginevalve actuation piston is vented to atmospheric pressure.

[0042] As the proportional valve moves from state 2 to state 3, the areathrough which high-pressure fluid from the high pressure rail 56 canflow into the control volume 26 above the engine valve actuation pistonincreases nonlinearly. (See FIG. 6). Similarly, as the proportionalvalve moves from state 2 to state 1, the area through which fluid canflow out of the control volume 26 above the engine valve actuationpiston to drain 37 increases nonlinearly (See FIG. 5). Thus the geometryof the proportional spool valve has been designed with regions of highand low gain. The low gain region provides fine control for take-off andseating velocities, while the high gain region provides the large flowarea required to achieve maximum engine valve velocities. Thisfacilitates more accurate control of the engine valve during seating andtake-off. These areas need more accuracy so that proper seatingvelocities and valve overlap are achieved throughout the full range ofengine speed and temperature.

[0043] To better describe the function of the exemplary hydraulicsystem, the following description traces the system through a completeengine valve operating cycle, mimicking results from a nodal hydraulicssimulation. The specific simulation model used 100° C. 0W30 syntheticmotor oil at an engine speed of 6000 rpm, though simulations at lowerengine speeds have also been run.

[0044] An exemplary valve event may be described as follows. Initiallythe supply pilot valve 22 is open and the vent pilot valve 20 is closed(as illustrated in FIG. 1). This keeps the proportional valve spool inthe venting (rightmost) position (State 3, FIGS. 5 & 6). Specifically,the flow area between engine valve actuation piston control volume 26and vent 37 is at a maximum (state 3, FIG. 6) and the area betweenengine valve actuation piston control volume and the high-pressure railis closed (state 3, FIG. 5). As a result, the engine valve is forcedclosed against its seat by the return spring 32.

[0045] To initiate valve opening, the supply pilot valve 22 is openedand the vent pilot valve 20 is closed. This allows fluid to flow fromthe control volume of the proportional spool valve to vent. As a result,the proportional spool begins to move from state 3. The vent pilot valve20 is left open long enough for the proportional spool to pass throughstate 2 and into state 1. However, the proportional valve is onlyallowed to travel until just a small flow area in the low gain region ofstate 1 is open between the high-pressure rail (FIG. 5) and engine valveactuation piston control volume 26. This results in a slow take-off ofthe engine valve. The speed of this take-off will vary depending onwhere the proportional valve is stopped. Then the vent pilot valve 20 isopened once again so that the proportional spool moves to a positionthat opens a larger flow area between the high pressure rail and theengine valve actuation piston's control volume. This results in a rapidopening of the engine valve after the initial slow takeoff.

[0046] The engine valve now must stop at the desired lift, in thisparticular example, 11 mm. To do this, the proportional spool will bemoved to state 2 in which the control volume above the engine valve ishydraulically locked. This is achieved by closing the vent pilot valve20 and opening the supply pilot valve 22 for the required amount oftime. The engine valve will stay in this position until it is commandedto return. At this point the kinetic energy in the engine valve is fullyconverted into potential energy of the fluid in the control volume andthe engine valve return spring. This trade off between kinetic andpotential energy occurs several times while the control volume ishydraulically locked, which can result in a slight oscillation of theengine valve position. To reduce this effect and to recover some of thekinetic energy in the proportional valve spool, a check valve could alsobe placed between the control volume 26 of the engine valve actuator andthe high-pressure rail 56 in order to damp out any high pressure spikesthat may occur during operation.

[0047] Next, the supply pilot valve 22 will be opened again long enoughto move the proportional valve to the high gain region of state 3, andthen closed, at least before vent pilot valve 20 is again opened. Toreiterate, at this point the flow area between the engine valve controlvolume 26 and vent 37 is a maximum. Therefore, the engine valve willaccelerate very quickly toward its seat via the stored energy in thevalve spring.

[0048] In order to seat the valve at the desired velocity, the flow areathat connects the engine valve control volume 26 and vent 37 must berestricted. This can be achieved by once again opening the vent pilotvalve 20 for a short period to reposition the proportional valve to alow gain in state 3. This seating velocity will change depending onwhere the proportional valve is stopped in this region.

[0049] This completes one engine valve cycle. In order to prepare thesystem for the next event, all components will be repositioned to theirinitial conditions. The only component that is out of position is theproportional valve. The supply pilot valve 22 is again opened, returningthe proportional valve to a position of maximum flow area in state 3.This reestablishes a reference point at the beginning of each valveevent, so that errors in proportional valve positioning do notaccumulate, one valve cycle to another. In this way, the seatingvelocities desired at different engine speeds, loads and temperaturescan be achieved by changing the position at which the proportional spooldwells. This can be facilitated further by varying the pressure in thelow-pressure rail if desired, thus accomplishing finer control of theproportional spool valve.

[0050] In a simulation of the system described above, an engine valveactuator consisting of two concentric pistons was used, as illustratedin FIG. 7. Instead of using one actuator with a relatively large areaexposed to pressure to drive the engine valve through its entire stroke,the large piston (boost piston 60) is used only initially to achievepeak velocities before reaching a mechanical stop, while the remainderof the stroke is accomplished using a smaller telescoping piston (drivepiston 62). Specifically, when the engine valve, particularly an exhaustvalve, initiates lift from its seat, in-cylinder pressure remainssubstantial. In addition, maximum engine valve acceleration is alsorequired at this time. As a result, a greater force is needed to actuatethe engine valve through the beginning of its stroke while a much lowerforce is required for the remainder of the stroke. The present inventionsystem does not rely on using the two concentric piston design, as itwill also function if just one actuator is used. However, the twoconcentric piston design requires less fluid from the high pressure railfor each valve cycle, and thus requires less energy for valve operation.

[0051] One can also use a hydraulically returned engine valve as opposedto a spring returned valve, as illustrated in FIG. 8. In thisembodiment, the valve actuator comprises a piston 64 having across-sectional area A₁ on a piston rod 66 having a cross-sectional areaA₂. Chamber 68 is permanently coupled to the high pressure rail, andchamber 70 is switchable by the proportional valve between the highpressure rail and the vent. Consequently, the maximum valve openingforce is equal to the pressure of the high pressure rail times A₂ andthe maximum valve closing force is equal to the pressure of the highpressure rail times A₁−A₂. Because of the functional relationshipbetween spring force and stroke, one can achieve essentially the samevalve dynamics as a hydraulically returned valve with a smaller diameteractuator.

[0052] With a return spring, the spring closing force is at a minimumwhen one desires a large opening force for maximum acceleration againstpeak cylinder pressure. With hydraulic return, the closing force of ahydraulically returned engine valve is constant and therefore will behigher than that of a spring when the valve is seated. Therefore, theforce characteristic of a mechanical spring is more desirable forreturning the engine valves than a single piston return mechanism.

[0053] Instead of using a closed center 3-way proportional valve, thehydraulically returned system can also be constructed using a closedcenter 4-way proportional valve (FIG. 9). Like the closed center 3-wayproportional valve, its position can also be infinitely variedthroughout 3 flow states.

[0054] State 1: The high pressure fluid is allowed to flow from the highpressure rail to a control volume 70 above the engine valve actuationpiston 64 while the fluid in chamber 68 acting below the engine valveactuation piston 64 is vented to tank.

[0055] State 2: The proportional valve is centered between its hardstops, trapping fluid in the control volume 70 above the engine valveactuation piston and in the control volume 68 below the engine valveactuation piston, thus creating a hydraulic lock.

[0056] State 3: The fluid in the control volume 70 above the enginevalve actuation piston is vented to atmospheric pressure while highpressure fluid is allowed to flow from the high pressure rail to thecontrol volume 68 below the engine valve actuation piston 64.

[0057] As the proportional valve moves from state 2 to state 1 the areathrough which high-pressure fluid can flow into the control volume 70above the engine valve actuation piston 64 increases nonlinearly(similar to FIGS. 5 & 6). At the same time the flow area between thefluid below the engine valve actuation piston 64 and tank increasesnonlinearly. Similarly, as the proportional valve moves from state 2 tostate 3, the area through which fluid can flow out of the control volume70 above the engine valve actuation piston 64 to tank increasesnonlinearly. At the same time the flow area between the fluid below theengine valve actuation piston 64 and the high-pressure rail increasesnonlinearly.

[0058] In all the systems described, the proportional valve useshydraulic force to oppose the pressure in its control volume.Alternatively, the proportional valve can also use a spring to supplypart or all of the opposing force.

[0059] Any of the 3-way proportional valve systems can take advantage ofthe recovery systems known in the art. The low-pressure rail used foractuating the 3 or 4 way proportional valve can be used for thelow-pressure source of the recovery system if that system isimplemented.

[0060] The present invention has many advantages for both diesel andgasoline engines, as well as similar engines powered with alternatefuels. These advantages include:

[0061] Infinitely variable engine valve timing for both opening andclosing times.

[0062] Infinitely variable engine valve lift from the valves seat to itsmaximum lift position.

[0063] Infinitely variable valve open and/or close time duration.

[0064] The proportional 3 or 4 way valve has low gain flow regions forfine control at valve take-off and seating. It also has high gain flowregions for maximum flow allowing increased speed of the engine valve sothat airflow into the engine cylinders can be maximized.

[0065] The system can allow the engine valve profile to benon-symmetric.

[0066] The system is capable of an infinitely varying the slew rate ofthe engine valve independent of rail pressure.

[0067] The system does not require a slow take-off and landing.Specifically, the valve can begin opening with maximum acceleration orseat at maximum velocity if desired.

[0068] The system does not need a lash adjustment system, specifically:

[0069] the system is unaffected and can compensate for the growth ofengine components (specifically valve train components) due to thermalexpansion,

[0070] the system is unaffected and can compensate for engine valverecession due to wear of the valve seat and the engine valve, and

[0071] the system is unaffected and can compensate for tolerance stackup between valve train components resulting from initial assembly andmanufacturing tolerances.

[0072] The system can compensate for varying working fluid viscosity dueto temperature, age, etc.

[0073] The system can optimize the amount time at which air is meteredinto the engine cylinder thus optimizing the combustion event at thefull spectrum of engine operating conditions resulting in:

[0074] maximum power,

[0075] lower emissions,

[0076] reduced emissions by controlling fuel/air mixing,

[0077] reduced heat rejection by reduction of unnecessary in cylinderair motion, and

[0078] high BMEP combustion schemes to improve catalyst light-off,reduce startup emissions.

[0079] The system can be operated in such a way that engine braking willresult, specifically by shutting off the injector during braking andopening the exhaust valve at the top of the compression stroke todissipate the compression energy.

[0080] The engine cycle can be varied to allow for:

[0081] 2 stroke operation.

[0082] Multiple stroke operation (such as, by way of example 2-stroke to4-stroke, 4-stroke to 6-stroke or 8-stroke operation, etc.) byeliminating one or more pairs of strokes from the normal engineoperating cycle, with the valves being controlled during these pairs ofstrokes for minimum energy loss and/or other considerations.

[0083] The system can allow for internal exhaust gas recirculation(EGR). As a result the EGR valve can be removed.

[0084] Variable compression ratio.

[0085] Miller cycle operation—Maximum cylinder pressure control withhigh expansion ratio for maximum thermodynamic efficiency.

[0086] Atkinson cycle operation.

[0087] Reduced heat rejection by reduction of unnecessary in cylinderair motion.

[0088] High BMEP combustion schemes to improve Catalyst light-off,reduce startup emissions.

[0089] Improved Cranking and Cold Start.

[0090] Reduced white smoke and diesel “fuel” smell during startup/coldtemperature idle/high altitude operation

[0091] High altitude compensation.

[0092] Variable torque curves to better fit duty/drive cycle of vehicle.

[0093] Increased torque at low speeds for better driveability, potentialvehicle fuel economy improvements.

[0094] The system will operate more efficiently with a sequentiallyapportioned pump.

[0095] The low-pressure rail can be replaced with an accumulator that issupplied by the return flow of the engine valve actuator.

[0096] Because the engine valve motion can be varied so that air can bethrottled at the engine valve, the throttle body can be eliminated.

[0097] Operation of the turbo can be optimized at all engine operatingconditions.

[0098] Cylinder deactivation for improved vehicle fuel economy.

[0099] This 2-stage system has the capability of satisfactorilycontrolling engine valves at very high engine speeds (from idle speedsto 10,000 RPM). In addition, the critical regions of valve take-off andseating can be controlled with accuracy and precision while providingthe features of infinitely variable valve timing, duration and lift. Thesystem also has the capability of significantly increasing the amount ofair that can be supplied to an engine's cylinders throughout the fullrange of engine speed by adjusting valve timing and duration to maximizethe dynamic effects of flow into and out of the combustion chamber atall engine speeds.

[0100] While an exemplary embodiment and various alternate embodimentsof the present invention have been disclosed herein, it will be obviousto those skilled in the art that various changes in form and detail maybe made therein without departing from the spirit and scope of theinvention.

What is claimed is:
 1. Apparatus for opening an engine valve comprising:a hydraulic actuator disposed with respect to the valve to encourage thevalve toward a valve open position by the pressure of a fluid in thehydraulic actuator; a proportional spool valve having a spoolhydraulically moveable between a first position coupling a source offluid under a first pressure to the hydraulic actuator and a secondposition coupling the hydraulic actuator to a reservoir of fluid under asecond pressure, the second pressure being less than the first pressure;electrically controlled valving hydraulically controlling the positionof the spool between the first and second positions; and, a valve returnreturning the valve to a closed position.
 2. The apparatus of claim 1wherein the spool valve has a third position between the first andsecond positions blocking the source of fluid under a first pressurefrom the hydraulic actuator blocking the hydraulic actuator from thereservoir of fluid under the second pressure.
 3. The apparatus of claim1 wherein the second pressure is atmospheric pressure.
 4. The apparatusof claim 1 wherein the spool is stepped in diameter to shape the areaversus spool position for flow between the source of fluid under thefirst pressure and the hydraulic actuator, and the hydraulic actuator tothe reservoir of fluid under the second pressure.
 5. The apparatus ofclaim 1 wherein the valving comprises double solenoid latching spoolvalves.
 6. A method of opening an engine valve comprising: providing ahydraulic actuator disposed with respect to the valve to encourage thevalve toward a valve open position by the pressure of a fluid in thehydraulic actuator; coupling the hydraulic actuator to a proportionalspool valve having a spool hydraulically moveable between a firstposition coupling a source of fluid under a first pressure to thehydraulic actuator and a second position coupling the hydraulic actuatorto a reservoir of fluid under a second pressure, the second pressurebeing less than the first pressure; and, hydraulically controlling theposition of the spool between the first and second positions byelectrically controlled valving.